Hydrocarbon fluid, ejector refrigeration system

ABSTRACT

A refrigeration system particularly suitable for an automobile air conditioning system is disclosed. The system includes an ejector to raise the pressure of the working fluid to the condenser pressure and a boiler which utilizes waste heat from the automobile engine. The boiler produces a near sonic velocity saturated vapor for input to the primary nozzle of the ejector. The near sonic velocity input maximizes the stagnation temperature and pressure of the fluid entering the primary nozzle for a given low or limited temperature heat source such as automobile waste heat. The refrigeration system uses a working fluid having a property such that its entropy when in a saturated vapor state decreases as pressure decreases, which maximizes the input pressure for a given heat source and prevents the vapor from condensing in the ejector. A high efficiency primary nozzle is also provided, allowing the system to take advantage of the decreasing entropy property of the saturated vapor. A refrigerant storage subsystem may be included in the system to provide an adequate transient response to changes in the waste heat output from the automobile engine and during the engine warm up period.

RELATED APPLICATIONS

This application is a continuation-in-part of copending application Ser.No. 07/889,615, filed May 27, 1992, now abandoned, which is acontinuation of Ser. No. 07/598,141, filed Oct. 16, 1990, now U.S. Pat.No. 5,117,648, issued Jun. 2, 1992.

FIELD OF THE INVENTION

This invention relates to the field of refrigeration and moreparticularly to automotive air conditioning systems.

BACKGROUND OF THE INVENTION

In the basic vapor-compression refrigeration cycle, the temperature of aworking fluid, or refrigerant, is reduced below the environmentaltemperature by an expansion process. Energy is then transferred as heatfrom the space to be cooled to the working fluid in an evaporator. Toreuse the refrigerant, it is repressurized to raise its temperatureabove the environmental temperature. Heat is then transferred from therefrigerant to the environment in a condenser. The cycle then repeats.

Chloroflourocarbons (CFCs) are often used as the working fluid invapor-compression refrigeration systems. However, in recent years, CFCpollution and particularly the destructive effect of CFCs on the earth'sozone layer have become of increasing concern. A major source of CFCpollution is the automobile air conditioning system, which typicallyuses the CFC freon as the working fluid. The CFCs are prone to leak outof the air conditioner both during the useful life of the airconditioner and when the vehicle is discarded. The state of Vermont andcities in California have even banned the use of automobile airconditioners which use CFCs as the working fluids.

Refrigerant systems have been devised in which the refrigerant is waterand which use a steam ejector rather than a compressor in the coolingcycle. Railroad passenger cars were cooled with a steam ejector systemduring the era of steam locomotion. This system was practical at thetime, since there was a ready source of high pressure steam and the sizeof the system was not critical. The steam ejector system was replaced byother types of air conditioning systems when diesel locomotives replacedsteam locomotives.

In the 1970s, a solar powered refrigeration system was proposed andanalyzed, although never constructed. In this system, solar energy wasused to boil a working fluid. The output from the boiler was a highvelocity, subsonic (Mach number less than 0.2) vapor which entered anejector system. Water was found to be the working fluid that gave thehighest coefficient of performance.

Water has several disadvantages as a working fluid, however,particularly for use in automobiles. Water is liable to freeze in winterin temperate climates, causing damage to the system. Also, water systemsrequire a low operating pressure in the evaporator and must be fairlylarge, so that they are difficult to fit into small vehicles.

Ejector refrigeration systems powered by automobile engine waste heathave been used. However, these prior art systems use a CFC refrigerantand do not account for the problems in the transient responsecharacteristics of the ejector system due to fluctuations in the wasteheat output of the engine, such as during engine start-up.

SUMMARY OF THE INVENTION

The refrigeration system of the present invention uses a hydrocarbon,such as isopentane, butane, or pentane, as the working fluid, ratherthan a CFC. The energy input to the working fluid may be the waste heat,such as from the coolant or exhaust gas of an automobile engine, whichis used to evaporate the working fluid in a boiler. The boiler producesa near sonic saturated vapor which enters an ejector.

The ejector replaces the compressor in the traditional vapor-compressionair conditioning system. In the ejector, the fluid from the evaporator,the secondary fluid, is pulled into the ejector and entrained into thefluid from the boiler, the primary fluid. The enthalpy of the primaryfluid from the boiler is converted into kinetic energy which is used toincrease the pressure of the entrained secondary fluid. The ejectorincludes a converging-diverging nozzle through which the primary workingfluid travels, a mixing region where the primary and secondary fluidsare mixed, and a diffuser. At the exit of the converging-divergingnozzle, the primary working fluid is traveling at a supersonic velocityand at a low pressure. The pressure at this location is equal to orslightly lower than the saturation pressure of the working fluid at thedesired evaporator temperature. The primary and secondary flows mix inthe mixing section and their pressures are increased through a standingshock wave at the entrance to the diffuser. The pressure of the workingfluid is further increased as it flows through the diffuser. The exitpressure of the diffuser equals the saturation pressure at the condenserdesign temperature.

The high pressure fluid enters the condenser where heat is transferredto the environment and the vapor is condensed to a saturated liquid. Thesaturated liquid is then divided into two flows. One flow passes throughan expansion valve and returns to the evaporator. The second flow ispumped to a higher pressure and returns to the boiler. The magnitude ofthe pump work is a negligible fraction of the cooling effect of thesystem since it involves pumping a liquid.

The near sonic velocity input maximizes the stagnation temperature andpressure of the fluid entering the ejector for a given low or limitedtemperature heat source, such as automobile waste heat. Also, theworking fluid is one having a property such that its entropy when in asaturated vapor state decreases as pressure decreases. In this manner, ahigh efficiency primary nozzle may be used without risk of condensingthe vapor as it passes through the nozzle and the ejector may beoperated at a high compression ratio, thereby increasing systemperformance.

The system includes a refrigerant storage subsystem to provide anadequate transient response to variations in the heat source, such as,for example, during warm up of an automobile engine or idling at a stoplight.

DESCRIPTION OF THE DRAWINGS

The invention will be more fully understood from the following detaileddescription taken in conjunction with the accompanying drawings inwhich:

FIG. 1 is a schematic diagram of the refrigeration system of the presentinvention;

FIG. 2 is a schematic diagram of an ejector for use in the presentinvention;

FIG. 3 is a temperature-entropy process diagram for the presentinvention with isopentane as the working fluid;

FIG. 4 is a schematic cross-sectional view of a boiler for use in thepresent invention;

FIG. 5 is a plan cross-sectional view of the boiler of FIG. 4; and

FIG. 6 is a schematic diagram of a further embodiment of therefrigeration system of the present invention.

DETAILED DESCRIPTION OF THE INVENTION

The cooling system of the present invention is shown generally at 10 inFIG. 1. The cycle comprises a first section 12 of a fluid flow path anda flow diverter 14 which diverts fluid flow on the path 12 into secondand third sections 16, 18 of the fluid flow path.

An expansion valve 20 and evaporator 22 are located on the third pathsection 18. A pump 24 and boiler 26 are located on the second pathsection 16. An ejector 30 is located at the outputs of the evaporator 22and boiler 26. A condenser 32 is located at the output of the ejector 30on the first path section 12.

A refrigerant fluid travels around the circuit shown in FIG. 1. Thefluid exits the condenser 32 as a saturated liquid. It enters the flowdiverter 14 and is diverted into the second and third flow paths 16,18.The fluid on the third path 18 passes through the expansion valve 20,where its temperature is decreased to the design temperature of theevaporator. The fluid then enters the evaporator 22. A cooling fluid,such as a water-ethylene glycol mixture, from the cooling coil passesaround the working fluid. Since the temperature of the working fluid isbelow the environmental temperature, heat is transferred from thecooling fluid to the working fluid, thereby cooling the cooling fluid.The cooling fluid returns to the cooling coil to cool the space to berefrigerated, such as the passenger compartment of an automobile. Theheated refrigerant from the evaporator then passes through control valve32 and enters the ejector 30.

The fluid on path 16 from the flow diverter 14 is pumped by the pump 24to the boiler 26. The fluid being pumped by the pump 24 is always aliquid, which is significantly easier to pump than a vapor andaccordingly reduces the pump work requirement of the system.

The boiler 26 is shown more fully in FIGS. 4 and 5. Preferably, theboiler is an evaporator section of a heat pipe or tube. The liquidrefrigerant coming from the pump 24 enters a porous capillary region 38,through which the flow of the liquid along the boiler tube is due tocapillary action. Surrounding the region 38 is a jacket 42 in the formof a coil winding around the region 38. Hot waste fluid from the engineenters the jacket 42 through a conduit 44, circulates through the jacket38, and exits through a conduit 46. The hot waste fluid in the jacketcauses the refrigerant liquid to evaporate, which enters the core region50 of the boiler tube as a saturated vapor. The velocity of the vaporincreases with distance along the boiler tube, since vapor iscontinuously added to the core, which remains fixed in size. Thesaturated vapor leaves the core at a high velocity through exit 52 andenters the ejector 30.

In an automobile, the boiler may serve as the automobile's radiator.Additionally, or alternatively, heat from the engine coolant or exhaustgases may be used. In an alternative embodiment of the boiler design,the boiler may be constructed from a series of tubes arranged inparallel. This arrangement increases the amount of surface area for heattransfer relative to the core region volume, which allows an increasedheat flow.

The ejector 30 is shown more fully in FIG. 2. The vapor from the boiler,known as the primary refrigerant, enters the ejector through a conduit62 which leads into a converging-diverging nozzle 64. The vaporrefrigerant from the evaporator, known as the secondary refrigerant,enters the ejector through an annular opening 60 which surrounds theconverging-diverging nozzle 64. In the ejector, the enthalpy of theprimary fluid is converted into kinetic energy which is used to increasethe pressure of the entrained secondary fluid.

Immediately following the nozzle 64 is a mixing region 66 in which thesecondary refrigerant and the primary refrigerant mix. This mixingregion may be of either a constant area type or a constant pressuretype. A constant pressure mixing process is illustrated in FIG. 2 and isdescribed below. When the primary fluid enters the mixing region 66, itis traveling at a supersonic velocity and at a low pressure. Thepressure at this location is equal to or slightly lower than thesaturation pressure of the working fluid at the desired evaporatortemperature. The mixing region comprises two sections. The first section68 is a converging section in which the primary and secondary fluids aremixed and the secondary liquid becomes entrained within the primaryvapor. The second section 70 is cylindrical and a standing shock wave isset up in this region. The pressure of the refrigerant is increased inthe shock wave region. Following the shock wave region 70 is thediffuser 72, in which the pressure is further increased.

The velocity of the primary working fluid entering the ejector should benear its sonic velocity to maximize the stagnation enthalpy and pressurePreferably, the Mach number of the working fluid should be greater than0.9. This yields the maximum compression in the ejector per unit of heatinput to the boiler. The required temperature of the heat source isminimized by having a saturated vapor enter the ejector.

To avoid condensing the primary fluid as it travels through theconverging-diverging nozzle, the fluid should be one whose entropy of asaturated vapor decreases as the pressure decreases. Severalhydrocarbons exhibit such behavior. Suitable fluids include isopentane,butane, pentane, heptane, hexane, isobutane, octane, Refrigerant 113,Refrigerant 114, and Refrigerant C-318.

FIG. 3 is a temperature-entropy process diagram for isopentane. At state1, the liquid on path 16 enters the boiler. It follows the line fromstate 1 to state 2 through the boiler. At state 2 it enters the primaryinlet of the ejector. State 3 is the exit of the primary nozzle 64. Atstate 3, the fluid is travelling at a high velocity and exits theprimary nozzle. The secondary fluid enters the ejector at state 4 andbegins mixing with the primary fluid at state 3. State 5 is the exit ofthe mixing section. Between states 5 and 6, the fluid is in the shockwave region of the mixing section. At state 6, it enters the diffuser.At state 7, it exits the diffuser. The fluid changes from a vapor, state7, to a saturated liquid, state 9, on the T-S diagram as it passesthrough the condenser. When the flow divides, the flow on path 18expands and enters the evaporator at state 8. The fluid on flow path 16enters the boiler at state 1.

The system of the present invention provides a number of advantages overprior art ejector-based refrigeration systems. Because the working fluidmust remain in the vapor phase through the ejector, prior art ejectorsystems using a working fluid such as water, R-11, R-12, and R-21, areinherently limited in their performance. For such working fluids, theentropy of the saturated vapor increases as the pressure decreases.Thus, to avoid entering the two-phase region shown on atemperature-entropy phase diagram as the fluid is expanded to a lowerpressure in the primary nozzle of the ejector, either the fluid must besuperheated before entering the ejector nozzle, or an inefficientprimary nozzle must be used in the ejector, or both. For a fixedtemperature heat source, if the fluid were heated to a superheated vaporprior to entering the ejector, the ejector inlet pressure would be lowerthan the ejector inlet pressure would be if the fluid entered theejector in the saturated vapor state. Typical efficiencies for readilycommercially available nozzles are approximately 85%.

The performance of the prior art ejector refrigeration system isadversely affected by these limitations, especially if the system is tobe powered by a waste heat or other low temperature source. Thecoefficient of performance, COP, of the system is strongly dependent onthe entrainment ratio and the condenser pressure. These values affectthe heat input to the boiler, which provides the primary fluid tooperate the ejector for increasing the pressure of the working fluidleaving the evaporator, P4, up to the condenser pressure, P8.

The entrainment ratio is the ratio of the secondary to primary mass flowrate. The greater the entrainment ratio, the greater the COP, becausethe cooling effect at the evaporator is increased for a greater massflow rate through the evaporator. The entrainment ratio is affected byejector efficiency. If the primary mass flow rate into the ejector wereminimized for a fixed entrainment ratio, the cooling effect would beachieved with less energy.

The condenser pressure, P8 on FIG. 3, is determined by the application.The ejector exit pressure, at state 7 on FIG. 3, must be equal to orgreater than the condenser pressure. The ejector exit pressure isdependent on the entrainment ratio, the primary nozzle efficiency, andthe entering pressure. The maximum ejector exit pressure is obtainedwhen the ejector inlet pressure is as large as possible and when a highefficiency primary nozzle is used. The ejector inlet conditions can alsobe stated in terms of the stagnation temperature and pressure of theworking fluid. The greater the pressure Pl or the stagnation pressure atthe ejector inlet, the greater the entrainment ratio can be for a fixedcondenser pressure. Any operating restriction which lowers the ejectorinlet pressure or its stagnation pressure or requires a low efficiencyprimary nozzle to avoid condensing the working fluid will decrease thecoefficient of performance of the refrigeration system.

In the system of the present invention, the performance of the ejectorrefrigeration system is improved through the use of a high efficiencyprimary nozzle and ejector and by increasing the stagnation temperatureand pressure of the working fluid entering the ejector. This is achievedfirst by using a hydrocarbon fluid with the characteristic that theentropy of the saturated vapor decreases as the pressure decreases. Aworking fluid with this characteristic can enter the primary nozzle ofthe ejector as a saturated vapor, state 2 on the temperature-entropydiagram of FIG. 3, without first being superheated. Entering in asaturated vapor state maximizes the inlet pressure to the ejector for alimited temperature heat source, such as waste heat from an engine. Thepressure P1 at state 2 is fixed, because the maximum temperature towhich the fluid can be heated in the boiler is fixed by the heat source.If the fluid were superheated, the maximum inlet pressure which could beachieved would necessarily be less than P1. Thus, for a given heatsource, the pressure of the fluid in the saturated vapor state is themaximum that can be achieved.

The stagnation temperature and pressure of the working fluid enteringthe primary nozzle are increased by using a boiler which produces a flowof high velocity saturated vapor. High velocity refers to a Mach numberof at least 0.85 and preferably in the range of 0.90 to 0.95. Increasingthe velocity increases the stagnation pressure entering the ejector,which increases the entrainment ratio for a fixed condenser pressureAlso, the boiler outlet is placed near and preferably adjacent to theprimary nozzle to minimize the pressure drop between the boiler and theejector. Additionally, while a Mach number greater than 0.95 ispossible, if the Mach number ranged too near 1.0, heat fluctuations tothe boiler might cause velocity fluctuations in the boiler which wouldchoke the flow and cause a shock wave within the boiler. If so, the flowinto the ejector would become subsonic and the device would not operateuntil the system reestablished itself.

The present invention also provides a primary nozzle having a highefficiency. The efficiency is at least 95% and preferably at least 98%.Thus, as shown in FIG. 3, flow through the nozzle, from state 2 to state3, is substantially isentropic. Due to the property of decreasingentropy of the saturated vapor with decreasing pressure, the fluid doesnot undesirably condense (enter the two-phase region on thetemperature-entropy diagram) when passing through such a nozzle. A fluidthat did not exhibit this property would condense by entering thetwo-phase region in a nozzle of this efficiency.

To achieve this efficiency, the converging section of the primary nozzleis designed to receive a high velocity flow from the boiler. Readilycommercially available prior art nozzles are designed for receiving lowvelocity flows, generally having a Mach number of approximately 0.1. Incontrast, the high efficiency nozzle of the present invention isdesigned for receiving a high velocity flow, with a Mach number of atleast 0.85. The nozzle converging section is determined by the includedangle, length, and configuration. In the present invention, the lengthof the converging section is generally shorter than the length of theconverging section in prior art nozzles. The length, L, of theconverging section is determined from the following equation: ##EQU1##where D_(i) =inlet diameter

D_(th) =throat diameter, and

Θ=included angle of converging section.

The inlet diameter D_(i) is a function of the Mach number: ##EQU2##where m=primary mass flow rate

c=sonic velocity, and

M=Mach number.

Thus, for a larger Mach number, the inlet diameter and consequently thelength of the converging section are less.

Further, for commercially available converging-diverging nozzlesdesigned for smaller Mach numbers, such as 0.1, the pressure at theinlet is approximately equal to the stagnation pressure, P_(o). In thenozzle of the present invention, designed for much greater Mach numbers,the stagnation pressure P_(o) is much greater than the inlet pressure.Thus, a greater entrainment ratio can be used to reach the samecondenser pressure, P8.

Additionally, the present invention preferably provides an ejector whichoperates at an entrainment ratio within 98% of an ideal ejector An idealejector is one containing isentropic nozzles and diffusers. In thismanner also, system performance is enhanced

A further embodiment of the present invention is shown in FIG. 6. Thisembodiment includes modifications to the system described in conjunctionwith FIG. 1 to improve the system's transient response characteristicsand performance when using a waste heat driven ejector The production ofwaste heat from an automobile engine (the flow rate of engine coolantand exhaust gases) is variable and fluctuates as a function of theengine speed. During periods such as engine warm up and idling, wasteheat production is low and may not be sufficient to provide adequatecooling. During periods when the engine speed is high, waste heatproduction may be more than is needed for adequate cooling.

As shown in FIG. 6, a refrigerant storage system includes a highpressure storage tank 82 located in a fluid path 83 in parallel withfluid path 18 and a low pressure storage tank 84 located off the fluidpath at the exit of the evaporator 22. The storage system is used duringperiods when the cooling load on the evaporator 22 is less than orgreater than the energy being removed from the cooling fluid by theevaporating refrigerant. Under the first condition, when the coolingload is less than the energy being removed from the cooling fluid, therefrigerant flow through the evaporator 22 is reduced by opening a valve86 connecting the low pressure tank 84 to draw stored refrigerant fromthe low pressure storage tank. The flow of low pressure refrigerant tothe ejector 30 remains constant during this operation. A valve 88located before the high pressure refrigerant storage tank 82 is openedto allow the high pressure tank to accumulate the excess refrigerantflow. This process continues until either storage tank is emptied orfilled.

The reverse procedure is implemented during engine start up or warm-upconditions or when there is a sudden reduction in the waste heat fromthe engine, such as during idling of the automobile at a stop light. Inthis mode of operation, the flow of refrigerant from the diverter valve14 to the evaporator 22 is insufficient to satisfy the cooling load.Accordingly, additional refrigerant flow to the evaporator 22 issupplied from the high pressure refrigerant tank 82 through expansionvalve 90. To prevent the evaporator 22 from being flooded, the valve 86to the low pressure tank is opened to collect the excess refrigerantflow from the evaporator.

Control of the refrigerant storage system may be effected automaticallyby, for example, a microprocessor and appropriate sensors, such astemperature sensors.

The refrigerant storage system may also be used to charge the highpressure refrigerant storage tank during the time when the engine isshut off and is still hot.

For long term storage, the temperature of the high pressure storage tankis equal to the ambient temperature and its pressure is less than thedesign pressure of the condenser. A concentric tank design may beprovided in which an inner tank is used for the high pressurerefrigerant storage and outer tank is used for the low pressurerefrigerant storage. With this configuration, heat transfer occursbetween the high and low pressure tanks. This heat transfer may be usedto maximize the pressure and amount of fluid stored in the high pressuretank when it is being charged. This configuration also provides theminimum operating pressure in the lower pressure storage tank while therefrigeration system is operated from the storage system.

A recuperator 92 may also be included in the system. The fluid from thecondenser 32 travels through control valve 94 on flow path 96 to therecuperator 92. The recuperator is located on the flow path after theexit from the evaporator 22. Fluid is returned from the recuperator onflow path 98 to the diverter 14. The recuperator is a heat exchangerwhich allows some of the energy from the condensing process to be usedto heat the refrigerant flow from the evaporator 22 to increase itstemperature and velocity upon entering the ejector 30. This increase intemperature and velocity of the secondary flow enables the primary flowto be reduced through the ejector while maintaining the same dischargepressure to the condenser. Thus, the system is able to compress a largeramount of secondary refrigerant per unit of heat input to the system.

The present refrigeration system can be manufactured as a hermeticallysealed unit and integrated with the existing connections to the heating,ventilating, and cooling system of an automobile to form the airconditioning system of the automobile, either when the automobile isinitially manufactured or as a retrofit to an existing automobile. Theretrofit may also be a replacement of an existing CFC-based airconditioner. The present invention has fewer moving parts than prior artvapor compression air conditioning systems typically used in automobilesand, accordingly, is more reliable than existing systems. In addition,by using automobile engine waste heat, fuel economy of the vehicle isincreased. A valve may also be provided to bypass the air conditioningsystem and direct the waste heat to the passenger compartment whennecessary for heating. The present refrigeration system may also besized for use in refrigerated surface transportation, such asrefrigerated trucks or city buses, and has potential applications incogeneration and solar energy systems.

Suitable connections may be provided for the engine coolant water andthe exhaust gases and a set of coolant connections to carry the coldworking fluid, such as a water-ethylene glycol mixture, to theenvironmental control system of the automobile. Heat pipes may beprovided to transfer the energy from the engine coolant and exhaustsystems to the boiler and from the evaporator to the cold working fluid.This type of packaging provides a double wall type of protection toprevent accidental discharge of the refrigerant fluid during anautomobile accident.

The invention is not to be limited by what has been particularly shownand described, except as indicated in the appended claims.

We claim:
 1. A refrigeration system comprising:a vapor ejector cycleincluding a working fluid having a property such that entropy of theworking fluid when in a saturated vapor state decreases as pressuredecreases, the vapor ejector cycle comprising:a condenser located on acommon fluid flow path; a diverter located downstream from the condenserfor diverting the working fluid into a primary fluid flow path and asecondary fluid flow path parallel to the primary fluid flow path; anevaporator located on the secondary fluid flow path; an expansion devicelocated on the secondary fluid flow path upstream of the evaporator; aboiler located on the primary fluid flow path parallel to the evaporatorfor boiling the working fluid, the boiler comprising an axiallyextending core region having a substantially constant cross sectionalarea and a porous capillary region surrounding the core region, the coreregion extending a length sufficient to produce a near sonic velocitysaturated vapor; and an ejector having an outlet in fluid communicationwith the inlet of the condenser and an inlet in fluid communication withthe outlet of the evaporator and the outlet of the boiler and in whichthe flows of the working fluid from the evaporator and the boiler aremixed and the pressure of the working fluid is increased to at least thepressure of the condenser, the ejector inlet, located downstream fromthe axially extending core region, including a primary nozzle locatedsufficiently close to the outlet of the boiler to minimize a pressuredrop between the boiler and the primary nozzle, the primary nozzle ofthe ejector including a converging section having an included angle andlength preselected to receive the working fluid from the boiler as anear sonic velocity saturated vapor.
 2. The refrigeration system ofclaim 1, wherein the working fluid is a hydrocarbon.
 3. Therefrigeration system of claim 2, wherein the hydrocarbon is isopentane.4. The refrigeration system of claim 2, wherein the hydrocarbon isbutane.
 5. The refrigeration system of claim 2, wherein the hydrocarbonis pentane.
 6. The refrigeration system of claim 2, wherein thehydrocarbon is heptane.
 7. The refrigeration system of claim 2, whereinthe hydrocarbon is hexane.
 8. The refrigeration system of claim 2,wherein the hydrocarbon is isobutane.
 9. The refrigeration system ofclaim 2, wherein the hydrocarbon is octane.
 10. The refrigeration systemof claim 2, wherein the hydrocarbon is Refrigerant
 114. 11. Therefrigeration system of claim 2, wherein the hydrocarbon is RefrigerantC-318.
 12. The refrigeration system of claim wherein the boilercomprises an evaporator section of a heat pipe.
 13. The refrigerationsystem of claim 1, wherein the axially extending core region is axiallyaligned with the primary nozzle of the ejector.
 14. The refrigerationsystem of claim 1, wherein the axially extending core region includes anoutlet located in direct fluid communication with the primary nozzle ofthe ejector.
 15. The refrigeration system of claim 1, wherein the nearsonic velocity saturated vapor from the boiler has a Mach number of atleast 0.85.
 16. The refrigeration system of claim 15, wherein the nearsonic velocity saturated vapor from the boiler has a Mach number between0.90 and 0.95.
 17. The refrigeration system of claim 1, wherein theprimary nozzle of the ejector has an efficiency of at least 95%.
 18. Therefrigeration system of claim 1, wherein the primary nozzle of theejector has an efficiency of at least 98%.
 19. The refrigeration systemof claim 1, wherein the primary nozzle of the ejector is sized tomaximize the inlet stagnation pressure.
 20. The refrigeration system ofclaim 1, further comprising a pump located in the second fluid flow pathupstream of the boiler to pump the working fluid to the boiler.
 21. Therefrigeration system of claim 1, wherein the ejector is configured tooperate at an entrainment ratio of within 98% of an ideal ejector. 22.An air conditioning system for an automobile comprising:a working fluidhaving a property such that entropy of the working fluid when in asaturated vapor state decreases as pressure decreases; a condenserlocated along a common fluid flow path for condensing the working fluidto a saturated liquid; a diverter located downstream of the condenserfor diverting the working fluid into a primary fluid flow path and asecondary fluid flow path parallel to the primary fluid flow path; anexpansion device located along the secondary fluid flow path fordecreasing the pressure of the working fluid; an evaporator locatedalong the secondary fluid flow path downstream of the expansion devicefor evaporating the working fluid; means located along the primary fluidflow path for boiling the working fluid to produce a near sonic velocitysaturated vapor; and an ejector located at a junction of the primary andsecondary flow paths, the ejector including means, located downstreamfrom the boiling means and sufficiently close to the boiling means tominimize a pressure drop between the boiling means and the ejector, forreceiving the working fluid from the boiling means in the near sonicvelocity saturated vapor state and for increasing the velocity of theworking fluid to a supersonic velocity, the ejector further including asection downstream from the receiving and velocity increasing means tomix the working fluid from the boiling means with the working fluid fromthe evaporator and to increase the pressure of the mixed working fluidto at least the pressure at the condenser.
 23. The air conditioningsystem of claim 22, wherein the working fluid is a hydrocarbon.
 24. Theair conditioning system of claim 22, wherein the hydrocarbon isisopentane.
 25. The air conditioning system of claim 23, wherein thehydrocarbon is butane.
 26. The refrigeration system of claim 23, whereinthe hydrocarbon is pentane.
 27. The air conditioning system of claim 22,wherein the hydrocarbon is isobutane.
 28. The air conditioning system ofclaim 22, wherein the boiling means comprises:an evaporator section of aheat pipe comprising a core, a jacket surrounding the core, and a poroussection between the jacket and the core through which fluid flows bycapillary action, the core having a length preselected to produce a nearsonic velocity saturated vapor; a source of a heating fluid connected tothe jacket to heat fluid in the porous section and in the core, wherebythe working fluid in the porous section enters the core by evaporation;an inlet for supplying the working fluid to the porous section; and anoutlet from the core for supplying the working fluid as a near sonicvelocity saturated vapor to the ejector.
 29. The air conditioning systemof claim 22, wherein the ejector comprises:a converging-diverging nozzlethrough which the working fluid from the boiling means travels and fromwhich the working fluid from the boiling means exits at a supersonicvelocity and reduced pressure; an annular input nozzle surrounding theconverging-diverging nozzle through which the working fluid from theevaporating means enters the ejector; a mixing section in which theworking fluid from the boiling means and the working fluid from theevaporator are mixed, the mixing section including a shock wave regionfor creating a standing shock wave to increase the pressure of theworking fluid; and a diffuser to further increase the pressure of theworking fluid such that the pressure of the working fluid at the exit ofthe diffuser is at least the saturation pressure of the condenser. 30.The refrigeration system of claim 29, wherein the converging-divergingnozzle of the ejector has an efficiency of at least 95%.
 31. Therefrigeration system of claim 29, wherein the converging-divergingnozzle of the ejector has an efficiency of at least 98%.
 32. The airconditioning system of claim 22, wherein the ejector entrains theworking fluid from the evaporator within the working fluid from theboiling means and converts the enthalpy of the working fluid from theboiling means to kinetic energy to increase the pressure of the workingfluid from the evaporating means.
 33. The air conditioning system ofclaim 22, further comprising a pump located in the primary flow pathupstream of the boiling means to pump the working fluid to the boilingmeans.
 34. The refrigeration system of claim 22, wherein the convergingsection of the converging-diverging nozzle is sized to maximize theinlet stagnation pressure.
 35. The air conditioning system of claim 22,wherein the boiling means further comprises means for supplying wasteheat from an automobile engine.
 36. The air conditioning system of claim35, wherein the waste heat supplying means supplies heat from an enginecoolant fluid.
 37. The air conditioning, system of claim 35, wherein thewaste heat supplying means supplies heat from exhaust gases from theengine.
 38. The air conditioning system of claim 22, wherein the boilingmeans comprises an automobile radiator.
 39. The refrigeration system ofclaim 22, wherein the near sonic velocity saturated vapor from theboiling means has a Mach number of at least 0.85.
 40. The refrigerationsystem of claim 39, wherein the near sonic velocity saturated vapor fromthe boiler has a Mach number between 0.90 and 0.95.
 41. An airconditioning system for an automobile comprising:a working fluid havinga property such that entropy of the working fluid when in a saturatedvapor state decreases as pressure decreases; a condenser located along acommon fluid flow path for condensing the working fluid to a saturatedliquid; a diverter located downstream of the condenser for diverting theworking fluid into a primary fluid flow path and a secondary fluid flowpath parallel to the primary fluid flow path; an expansion devicelocated along the secondary fluid flow path for decreasing the pressureof the working fluid; an evaporator located along the secondary fluidflow path downstream of the expansion device for evaporating the workingfluid; a boiler located on the primary fluid flow path parallel to theevaporator for boiling the working fluid, the boiler comprising anaxially extending core region having a substantially constant crosssectional area and a predetermined length, and a porous capillary regionsurrounding the core region and cooperating with the core region todefine means to produce a near sonic velocity saturated vapor; and anejector located at a junction of the primary and secondary flow pathssufficiently close to the outlet of the boiler to receive the near sonicvelocity saturated vapor from the boiler, the ejector comprising:aconverging-diverging nozzle through which the working fluid from theboiler travels and from which the working fluid from the boiler exits ata supersonic velocity and reduced pressure, the converging sectionhaving an included angle and length preselected to receive the workingfluid from the boiler as a near sonic velocity saturated vapor; anannular input nozzle surrounding the converging-diverging nozzle throughwhich the working fluid from the evaporator means enters the ejector; amixing section in which the working fluid from the boiler and theworking fluid from the evaporator are mixed, the mixing sectionincluding a shock wave region for creating a standing shock wave toincrease the pressure of the working fluid; and a diffuser to furtherincrease the pressure of the working fluid such that the pressure of theworking fluid at the exist of the diffuser is at least the saturationpressure of the condenser.
 42. The air conditioning system of claim 41,wherein the working fluid is a hydrocarbon.
 43. The air conditioningsystem of claim 42, wherein the hydrocarbon comprises isopentane,butane, or pentane.
 44. The air conditioning system of claim 41, whereinthe converging-diverging nozzle of the ejector has an efficiency of atleast 95%.
 45. The air conditioning system of claim 41, wherein theconverging-diverging nozzle of the ejector has an efficiency of at least98%.
 46. The air conditioning system of claim 41, wherein the near sonicvelocity saturated vapor produced by the boiler has a Mach number of atleast 0.85.
 47. The air conditioning system of claim 42, wherein thenear sonic velocity saturated vapor produced by the boiler has a Machnumber between 0.90 and 0.95.